Disk drive systems have been used in computers and other electronic devices for many years for storage of digital information. Information is recorded on concentric memory tracks of a magnetic disk medium, the actual information being stored in the form of magnetic transitions within the medium. The disks themselves are rotatably mounted on a shaft or “spindle”, the information being accessed by means of transducers located on a pivoting arm, which moves radially over the surface of the disk. The read/write heads or transducers must be accurately aligned with the storage tracks on the disk to ensure proper reading and writing of information; thus the disks must be rotationally stable.
Electric spindle motors are used to rotate the discs in disk drive systems. Such spindle motors may have either a fixed shaft and a rotating sleeve or a fixed sleeve and a rotating shaft. In recent years, there was an increase in demand for smaller size and lighter weight spindle motors. There was also an increase in demand for higher density of memory capacity in data recording devices such as magnetic disks and optical disks used in computers. These developments led to an increased demand for technologies increasing motors' rpm speed and improving rotation accuracies in spindle motors used to rotate such disks.
To address this demand with respect to bearings used to support rotating shafts in data storage devices, there has been an increasing trend away from conventional ball bearings toward the adoption of fluid dynamic pressure bearings. Fluid dynamic pressure bearings support a rotating shaft by generating a fluid dynamic pressure within lubricating fluid, for example oil or air, when the shaft is rotated.
Fluid dynamic pressure bearings are well known in the art. Structures which employ fluid dynamic pressure bearings as bearings for spindle motor rotating shafts are also well known (see, for example, Japanese Patent No. 2937833). An example of a conventionally known spindle motor having a fluid dynamic pressure bearing is shown as the conventional example in FIG. 1.
As shown in FIG. 1, spindle motor 10 includes rotor assembly 14 and stator assembly 16. Rotor assembly 14 comprises rotating hub 20 supported by bearing sleeve 32. Yoke 22 is provided in the lower portion of hub 20 for supporting permanent magnet 24. Stator assembly 16 comprises stationary frame 18 fixedly attached to shaft 31. Shaft 31 is inserted into an inner cylindrical bore of bearing sleeve 32. Stator core 26 having winding 28 is secured to an inner portion of frame 18 such that the stator core with the winding are positioned in an opposing relationship with permanent magnet 24.
Rotating bearing sleeve 32 and shaft 31 supporting the bearing sleeve for rotation define bearing portion 30, more particularly shown in FIGS. 2b and 3b. Bearing portion 30 includes bearing sleeve 32 secured to the hub and bearing member 33 secured to the shaft 31. Bearing member 33 is partly placed inside a conically-shaped recess formed within bearing sleeve 32. Bearing gap 55 is formed in the conical shaped area between the corresponding and opposing surfaces of bearing member 33 on one hand and bearing sleeve 32 on the other hand. Lubricating reservoir 35 is formed between the protruding part of bearing member 33 and shield 39. In order to equalize pressure differences along the bearing gap communication path 34 connects the inner end with the outer end of bearing gap and with the lubricating reservoir. Bearing gap 55 and lubricating reservoir 35 are filled with lubricating oil 12. At least one fluid dynamic pressure-generating groove (not shown) can be provided on one or more of the above mentioned surfaces.
When bearing sleeve 32 rotates, the fluid dynamic pressure generated by the fluid dynamic pressure-generating groove in the radial and thrust directions, enables the rotating bearing sleeve to rotate in a suspended state around shaft 31, with a film of lubricating oil interposed therebetween.
During operation of the above described spindle motor, lubricating oil 12 enclosed in bearing gap 55 between bearing member 33 and bearing sleeve 32 ascends to the opening surface at the top edge portion of sleeve 32. This oil ascending phenomenon may be caused by volumetric changes from temperature change-induced expansion and contraction of the lubricating oil, expansion displacement of the bearing dimensions, internal movement caused by the pumping effect at the start and stop of sleeve's rotation or effects of centrifugal forces and dynamic pressure during rotation, and mainly due to external shock.
This type of ascending of the lubricating oil such that it reaches and overflows the opening surface of the bearing sleeve creates the problem of lubricating oil leakage. Leakage and depletion of the lubricating oil from the bearing sleeve results in insufficient fluid dynamic pressure, reduced lubrication, and, in some cases, burning through contact between the rotating shaft and the bearing sleeve. At the same time leaking lubricating oil can lead to head sticktion or head crash thus making the magnetic disc unreadable.
As shown in FIGS. 4(a)-(c), a gap widening portion 37 having a tapered surface 36 is provided in conventional fluid dynamic bearing structures at the upper portion of the bearing sleeve to prevent leakage of lubricating oil. Gap widening portion 37 gradually expands at a specified angle of inclination a, as measured between the inner surface of bearing sleeve 32 and the axis of the shaft at the gap opening edge area. Thus, the upper portion of the gap gradually widens in the direction of the opening surface. Further, as shown in FIG. 4(c), the bearing may also include a lubricating oil reservoir 38 disposed on the inner surface of bearing sleeve 32, specifically, on the inside of tapered surface 36.
As disclosed in Japanese Patent No. 2937833, an oil collecting groove may be disposed on the inner surface of the bearing sleeve. A gap changing portion is also provided in the disclosed construction, wherein the gap expands towards the opening surface of the bearing sleeve. Taking a as the angle of gap's expansion towards the outside, an inner surface of the gap changing portion may be inclined at the angle α of 0° or greater. As is disclosed in the '833 patent, a gap inclination angle α of 0° indicates that it is acceptable to have a partial area of the gap changing portion being parallel to the rotating shaft.
As shown in FIGS. 2b and 3b, shield 39 is placed at each end of shaft 31 to seal the opening surface of bearing sleeve 32 from the environment, thus preventing lubricating oil from splashing onto the magnetic disk. However, shield 39 does not prevent oil from ascending the shaft. Typically, to prevent oil from splashing onto the shaft and then ascending the shaft into the environment, separate washer 40 is placed onto the shaft to trap the splashing and ascending fluid. As shown in FIG. 3c, washer 40 is provided with a pair of sparings 42 allowing for the oil to be injected into the bearing gap through oil filling holes 41 in shield 39. Accordingly, sparings 42 have to be aligned with oil filling holes 41. Therefore, washer 40 has to be installed with very exact tolerances and needs to be coated with an oil repelling agent before installation in order to ensure that lubricating oil is conveyed into the bearing gap. The oil repelling agent typically consists of a material with very low surface tension, for example fluorocarbon compounds. These processes of coating and installing washer 40 are expensive, time consuming and difficult to accomplish efficiently.